Gear mechanism of power transmitting system

ABSTRACT

The present invention provides a gear mechanism of a power transmitting system, which includes a crank gear operatively coupled with a crankshaft, and a first driven gear that is coupled with a first balance shaft via a damping mechanism that allows relative rotation between the driven gear and the balance shaft. The damping mechanism includes a friction damper that generates frictional force when the first driven gear and a counter gear fixed to the first balance shaft rotate relative to each other within a predetermined rotational phase range or angle, and a stopper rubber or rubbers that elastically deforms to generate elastic force when the first driven gear and the first balance shaft rotate relative to each other beyond the predetermined rotational angle.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to a gear mechanism of a powertransmitting system that is favorably used as a balancer apparatus of aninternal combustion engine.

2. Discussion of Related Art

As well known in the art, in a balancer apparatus of an internalcombustion engine, a balance shaft provided with an unbalance weight isoperatively coupled with a crankshaft via a gear mechanism, wherebyrotational force of the crankshaft is transmitted to the balance shaft.In the balancer apparatus, the balance shaft rotates in synchronizationwith the crankshaft, whereby inertial force generated by reciprocationof an engine piston is cancelled, and vibration of the engine isaccordingly reduced.

Since explosive combustion in the internal combustion engine takes placeintermittently, the magnitude of the rotational force transmitted fromthe crankshaft to the balance shaft is not constant or fixed, but ratheris always fluctuating.

The inventors have confirmed that, among frequency components includedin the fluctuations of the rotational force, a secondary component of afundamental frequency that results from engine combustion occurring oncein every two rotations of the crankshaft, and a sextic component that isamplified by torsional resonance of the crankshaft are relatively largecompared to a component (primary component) of the fundamental frequencythat is determined according to the speed of rotation of the crankshaft.

The balancer apparatus receives the rotational force including thevibration components of different frequencies as described above, andtherefore vibration occurs in the gear mechanism, in particular, in ameshing portion(s) of the gears. Such vibration may result in generationof noise and reduction in the durability of the gears.

Thus, a balancer apparatus has been proposed wherein a damping mechanismformed by, for example, a spring or springs is inserted in arotational-force transmission path from the crankshaft to the balanceshaft so as to damp the vibration components of the rotational force.

In order to effectively damp a high-frequency component of thefluctuations in the rotational force, such as the sextic component ofthe fundamental frequency, by using the damping mechanism, the springconstant of the spring(s) must be set to a sufficiently low value so asto reduce the natural frequency of a vibration system formed by thebalancer apparatus. However, if the spring constant is merely set to alow value, the spring(s) may be excessively deformed in response to arapid increase in the rotational force transmitted from the crankshaftupon, for example, acceleration of the engine. Thus, the dampingmechanism may be damaged due to the deformation. Moreover,characteristics of the spring may be substantially lost by so-calledbottoming or the like, whereby the damping mechanism may cease tofunction properly.

In view of the above situation, a balancer apparatus in which a dampingmechanism provides non-linear spring characteristics has been proposedin, for example, Japanese Laid-Open Patent Publication No. 60-192145.

FIG. 22 shows a cross-sectional structure of a main part of one exampleof the balancer apparatus. As shown in FIG. 22, the balancer apparatusincludes a rotary shaft 100 operatively coupled with a balance shaft(not shown), and a generally cylindrical gear 110 that surrounds theouter periphery of the rotary shaft 100 and operatively coupled with acrankshaft (not shown). The rotary shaft 100 has a plurality of radiallyprotruding driving pieces 102 formed on its outer periphery. The gear110 also has a plurality of radially protruding driving pieces 112formed on its inner periphery so as to be located between thecorresponding driving pieces 102 of the rotary shaft 100.

Damper chambers 120 are formed between the respective driving pieces 102of the rotary shaft 100 and the corresponding driving pieces 112 of thegear 110, and an elastic member 130 is disposed in each damper chamber120. Moreover, clearances 132 are formed between each elastic member 130and the corresponding driving pieces 102 and 112. In the balancerapparatus thus constructed, the driving pieces 102 and 112 and theelastic members 130 form the damping mechanism.

The operation of the damping mechanism will be now described. As therotary shaft 100 rotates relative to the gear 110, the clearances 132are reduced, and the driving pieces 102 and 112 then abut on therespective elastic members 130. As the rotary shaft 100 further rotatesrelative to the gear 110, the elastic members 130 are elasticallydeformed, thereby generating elastic force according to the amount ofrelative rotation. This elastic force (more specifically, torque basedon this elastic force) acts against the relative rotation between therotary shaft 100 and the gear 110.

Referring to FIG. 23, the solid line indicates the relationship betweenthe angle θr of the relative rotation between the rotary shaft 100 andthe gear 110 and the elastic force (torque) T. The two-dot chain lineindicates the relationship between the relative rotational angle θr andthe elastic force T in a comparative example. In the comparativeexample, the clearances 132 are not formed, and the natural frequency ofthe vibration system is reduced merely by setting the spring constant ofthe elastic members 130 to a low value.

As indicated by the solid line of FIG. 23, when the relative rotationalangle θr is within a predetermined rotational phase range or angle(θr<θ1), the elastic members 130 are not elastically deformed, wherebythe elastic force T is “zero”. Thus, by forming the clearances 132between each driving piece 102, 112 and the corresponding elasticmembers 130 so as to provide a relative rotational phase range in whichthe elastic force T is not produced, the natural frequency of thevibration system formed by the balancer apparatus can be reduced withoutsignificantly reducing the spring constant of the elastic members 130.

When the rotary shaft 100 and the gear 110 rotate relative to each otherbeyond the predetermined rotational angle (θr>θ1), the elastic force Tincreases with the relative rotational angle θr. As compared with thecomparative example, the relative rotational angle θr is limited to arelatively small value even when the elastic force T becomes extremelylarge (T=Tmax), that is, when the rotational force transmitted from thecrankshaft to the balancer apparatus becomes extremely large(θmax1<θmax2). Thus, the elastic members 130 are not excessivelydeformed.

Thus, according to the balancer apparatus, a high-frequency component ofthe fluctuation in rotational force can be damped without causing anydamage and deterioration in the function of the damping mechanism whenthe rotational force from the crankshaft rapidly increases upon, forexample, acceleration of the engine.

Such a damping mechanism having a non-linear spring characteristic cancertainly reduce the natural frequency of the vibration system formed bythe balancer apparatus, and damp the high-frequency component of thefluctuations in the rotational force, while avoiding any damage anddeterioration in the function of the damping mechanism.

However, the reduction in the natural frequency of the vibration systemmay cause a problem as follows: the natural frequency is reduced to beequal to a frequency that is close to that of a low-frequency component,such as the secondary component of the fundamental frequency of theengine, which is included in the fluctuations in the rotational force.As a result, a resonance phenomenon occurs in the balancer apparatus dueto the low-frequency component of the fluctuations in rotational force.Thus, vibration resulting from the resonance phenomenon cannot beprevented.

The aforementioned problem occurs not only in the above-describedbalancer apparatus of the internal combustion engine, but generallyoccurs in a gear mechanism of a power transmitting system that transmitsrotational force by using gears.

SUMMARY OF THE INVENTION

The present invention has been developed in the light of the abovesituations. It is therefore an object of the present invention toprovide a gear mechanism of a power transmitting system that is capableof favorably preventing or reducing the occurrence of a resonancephenomenon due to high-frequency and low-frequency components.

To accomplish the above object, the present invention provides forexample a gear mechanism of a power transmitting system comprising firstand second rotating members disposed coaxially with each other, and adamping mechanism interposed therebetween, wherein the damping mechanismincludes a damping member that generates damping force for limitingrelative rotation between the first and second rotating members, and atleast one elastic member that elastically deforms mainly when an angleof relative rotation between the first and second rotating membersexceeds a predetermined rotational angle, so as to apply elastic forceonto the rotating members in a direction opposite to that of therelative rotation.

The gear mechanism as described above may be applied to a constructioncomprising a first gear operatively coupled with a first rotary shaftand meshing with a second gear provided on a second rotary shaft,wherein the second gear forms one of the rotating members and the secondrotary shaft forms the other rotating member.

With the gear mechanism constructed as described above, the elasticforce of the elastic member is generated mainly when the two coaxiallydisposed rotating members (the second gear and the second rotary shaftin the above application) rotate relative to each other beyond thepredetermined rotational angle. Therefore, the natural frequency of avibration system including the gear mechanism can be reduced withoutsignificant reduction in the spring constant of the elastic member. As aresult, a high-frequency component of fluctuations in rotational forcecan be damped while avoiding any damage and degradation in the functionof the damping mechanism. When the two coaxially disposed rotatingmembers rotate relative to each other within the predeterminedrotational phase range or angle, on the other hand, the damping membergenerates damping force for limiting the relative rotation between therotating members. Therefore, the damping capability of the dampingmechanism can be enhanced, and a low-frequency component of thefluctuations in the rotational force can also be damped.

Thus, even where the rotational force including both low-frequency andhigh-frequency components as vibration components is transmitted to thegear mechanism constructed according to the present invention,occurrence of a resonance phenomenon due to the low-frequency andhigh-frequency components can be advantageously prevented withoutcausing any damage or degradation in the function of the dampingmechanism.

In one preferred form of the present invention, the damping membercomprises a friction damping member that is located between the tworotating members, so as to generate the damping force in the form offriction force that arises due to relative rotation between the tworotating members.

With the gear mechanism constructed as described above, the dampingforce does not significantly change with a change in the speed at whichthe two rotating members rotate relative to each other, and may be heldsubstantially constant. Therefore, the capability of damping, inparticular, a low-frequency component of the fluctuations in rotationalforce can be improved as compared with a structure that uses a so-calledoil damper or the like as the damping member. As a result, occurrence ofa resonance phenomenon due to the low-frequency component can be furtheradvantageously suppressed or prevented.

In a further preferred form of the present invention, each of the atleast one elastic member comprises a main deformation portion arrangedat an acceleration side of the elastic member that elastically deformsmainly when the two rotating members rotate relative to each otherbeyond the predetermined rotational angle as the rotational forcetransmitted between the two rotating members increases or is keptsubstantially constant, and a sub deformation portion arranged at adeceleration side of the elastic member that elastically deforms mainlywhen the two rotating members rotate relative to each other beyond thepredetermined rotational angle as the rotational force transmittedbetween the two rotating members decreases, the acceleration-sideelastic portion having a greater limit to elastic deformation thereofthan that of the deceleration-side elastic portion.

With the gear structure constructed as described above, theacceleration-side elastic portion elastically deforms when the tworotating members rotate relative to each other beyond the predeterminedrotational angle not only in the case where the rotational forceincreases but also in the case where the rotational force is keptsubstantially constant. Thus, the acceleration-side elastic portionfunctions to transmit the rotational force between the two rotatingmembers. Therefore, the acceleration-side elastic portion is morefrequently subjected to elastic deformation than the deceleration-sideelastic portion.

In view of the above, the damping mechanism of the invention may beconstructed such that the main deformation portion has a greater limitto elastic deformation thereof than that of the sub deformation portion.Thus, the main deformation portion is allowed to elastically deform to agreater extent, thus assuring improved durability of the elasticmember(s).

In the above preferred form of the invention, the main deformationportion and the sub deformation portion may be formed of a rubbermaterial, and the main deformation portion may have an elasticallydeformable portion whose volume is larger than that of an elasticallydeformable portion of the sub deformation portion.

At least one of the two rotating members may comprise a gear meshingwith a respective counter gear.

In the gear mechanism of the present invention, at least one of the gearand the counter gear may be a resin gear whose teeth comprise a resinmaterial.

In the above structure, impact that acts on a meshing portion betweenthe meshing gears is absorbed, and gear meshing noise can be thusreduced. Furthermore, the fluctuations in the rotational forcetransmitted between the gears, in particular, its high-frequencycomponent, can be advantageously damped.

Also, in the case where the other of the meshing gears that meshes withthe resin gear is a metal gear, the gear meshing noise can be reducedeven if the working accuracy of a tooth surface of the metal gear isrelatively low. This makes it possible to eliminate some process steps,such as shaving and polishing the tooth surface of the metal gear, andbacklash control by selection and adjustment of a shim commonly used forforming metal gears. Moreover, since resin gears are respectively meshedwith metal gears in the gear mechanism, disadvantages such as thermaladhesion between the gears can be avoided.

In the gear mechanism as described above, one of the meshing gears maybe a resin gear whose teeth are formed of a resin material, while theother of the meshing gears may be a metal gear whose teeth are formed ofa metal, and the resin gear may have a tooth width that is larger thanthat of the metal gear.

In the case where the resin gear and the metal gear mesh with eachother, respective tooth positions of the gears may be displaced fromeach other in the tooth-width direction of the gears due to an error inmounting the gears, vibration during rotation, and the like. In such acase, only a local portion of the tooth surface of the resin gear abutson the tooth surface of the metal gear, resulting in so-called localabutment. Since the resin gear is generally less wear-resistant and lessdurable than the metal gear, the resin gear may further be worn and/ordamaged as a result of the local abutment.

In view of the above point, the gear mechanism of the present inventionmay be constructed such that the resin gear has a tooth width that islarger than that of the metal gear. Therefore, even if the respectivetooth positions of the gears are displaced from each other, thedisplacement is covered, and abutment of local portions of the resin andmetal gears is avoided. As a result, otherwise possible wear and damageof the resin gear resulting from the displacement can be suppressed orprevented.

In the gear mechanism as described above, at least one elastic membermay be provided on one of the two rotating members, and at least oneabutting member may be provided on the other of the two rotatingmembers, each abutting member abutting on the corresponding elasticmember to cause elastic deformation thereof when the two rotatingmembers rotate relative to each other beyond the predeterminedrotational angle. In this arrangement, the strength of each abuttingmember as measured upon breakage of the abutting member due to elasticforce of the corresponding elastic member acting thereon may be set tobe smaller than the strength of a toothed portion of one of the meshinggears that is formed as the resin gear.

In the case where at least one of the gears is a resin gear, a toothedportion of the resin gear may be broken when it receives excessiverotational force since the strength of the resin gear is lower than thatof a metal gear. This may result in a problem such as biting of thegears.

In this respect, the gear mechanism of the invention may be constructedsuch that the abutting member is broken prior to breakage of the toothedportion of the resin gear, whereby mechanical coupling force between thetwo rotating members is rapidly reduced. With this arrangement, thebreakage of the toothed portion of the resin gear upon receipt of theexcessive rotational force is avoided, whereby the problem such asbiting of the gears can be prevented in advance.

In another preferred form of the present invention, the dampingmechanism comprises a plurality of elastic members as theabove-indicated at least one elastic member, each of which is providedon one of the two rotating members, and a plurality of abutting memberscorresponding to the respective elastic members are provided on theother of the two rotating members, for abutting on the correspondingelastic members to cause elastic deformation thereof when the tworotating members rotate relative to each other beyond the predeterminedrotational angle. Furthermore, the elastic members are located withrespect to the one of the two rotating members such that differentangles of relative rotation between the two rotating members are formedwhen the respective elastic members successively abut on thecorresponding abutting members.

With the above arrangement, the elastic members as a whole exhibit aneven more non-linear elastic characteristic when the two rotatingmembers rotate relative to each other. Therefore, the natural frequencyof the vibration system including the gear mechanism is dispersed ordiversified into a plurality of frequencies, whereby the dampingcapability of the damping mechanism is further enhanced. As a result,occurrence of the resonance phenomenon can be suppressed in a furtherpreferable manner.

In a further preferred form of the invention, the damping mechanismcomprises a plurality of elastic members, each of which is provided onone of the rotating members, and a plurality of abutting memberscorresponding to the respective elastic members, each of which isprovided on the other of the rotating members for abutting on thecorresponding elastic members to cause elastic deformation thereof whenthe rotating members rotate relative to each other beyond thepredetermined relative rotational angle, wherein the elastic members andthe abutting members are located with respect to the rotating memberssuch that the elastic members and the abutting members are spaced fromeach other at equal intervals in a direction of rotation of the rotatingmembers, and wherein the number of teeth of the gear being coupled bythe gear mechanism is set to an integral multiple of the number of theelastic members.

Thus an increased degree of freedom with which the gear is mounted maybe achieved.

In a further preferred form of the invention, the predetermined relativerotational angle is defined by the sum of angles by which each abuttingmember is spaced from corresponding end faces of the above-indicated atleast one elastic member which face the abutting member, as viewed in adirection of rotation of the rotating members.

The present invention may be applied to an internal combustion enginecomprising a crankshaft, at least one balance shaft and a gear mechanismaccording to the invention, wherein the at least one balance shaft isdriven by a rotational torque of the crankshaft.

In a further preferred form of the invention, the gear mechanism isarranged at the first balance shaft and comprises a driven gear beingdisposed on the first balance shaft and rotatable relative thereto, andwherein the driven gear is driven by a crank gear being fixedly securedon the crankshaft.

In a still further preferred form of the invention, the gear mechanismis arranged at the crank shaft and comprises a crank gear being disposedon the crankshaft and rotatable relative thereto, and wherein the crankgear drives a driven gear being fixedly secured on the first balanceshaft.

The internal combustion engine may comprise a second balance shaft beingoperatively coupled with the first balance shaft.

In a further form of the invention, the second balance shaft may bedriven by the crankshaft via the crank gear, an intermediate gear beingdisposed on an intermediate shaft and meshing with the crank gear, adriven gear being disposed on the second balance shaft and rotatablerelative thereto and meshing with the intermediate gear, and anadditional gear mechanism connecting between the driven gear and thesecond balance shaft.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a side view schematically showing the construction of a gearmechanism according to the first embodiment of the present invention,which gear mechanism is employed in a balancer apparatus of an internalcombustion engine.

FIG. 2 is a perspective view showing the meshing relationship of gearsin the gear mechanism of the first embodiment of the present invention.

FIG. 3 is a cross-sectional view showing the construction of a dampingmechanism of the gear mechanism of the first embodiment.

FIG. 4 is a cross-sectional view taken along line 4—4 of FIG. 3.

FIG. 5 is a schematic diagram showing the gears meshing with each otherin the gear mechanism of the first embodiment.

FIG. 6 is a diagram modelling the gear mechanism of the firstembodiment.

FIG. 7 is a graph showing a characteristic of fluctuations in theangular acceleration of a balance shaft in relation to the speed ofrevolution of the internal combustion engine.

FIG. 8 is a graph showing a characteristic of fluctuations in theangular velocity of a balance shaft in relation to the speed ofrevolution of the internal combustion engine.

FIG. 9 is a graph showing a characteristic of fluctuations in theangular velocity of a balance shaft in relation to the speed ofrevolution of the internal combustion engine.

FIG. 10 is a cross-sectional view showing the construction of a dampingmechanism in a gear mechanism according to the second embodiment of thepresent invention.

FIGS. 11A, 11B and 11C are schematic diagrams illustrating the meshingor engaging states of a resin gear and a metal gear.

FIG. 12 is a graph showing the relationship between the ratio of thetooth width of the resin gear to the tooth width of the metal gear andthe strength of a toothed portion of the resin gear.

FIGS. 13A and 13B are diagrams illustrating the manner of measuring thestrength of the toothed portion of the resin gear and the manner ofmeasuring the breakage strength of each projection provided in the resingear, respectively.

FIG. 14 is a cross-sectional view showing the construction of a dampingmechanism in a gear mechanism according to the fifth embodiment of thepresent invention.

FIG. 15 is a cross-sectional view showing the construction of a dampingmechanism in a gear mechanism according to the sixth embodiment of thepresent invention;

FIG. 16 is a cross-sectional view taken along line 16—16 of FIG. 15.

FIG. 17 is a cross-sectional view showing the construction of a dampingmechanism in a modified example of the gear mechanism of the sixthembodiment.

FIG. 18 is a cross-sectional view showing the construction of a dampingmechanism in a modified example of the gear mechanisms of the firstthrough fifth embodiments.

FIGS. 19A and 19B are schematic diagrams useful for explaining anengaging state of teeth of a resin gear with teeth of a metal gear.

FIG. 20 is a schematic diagram showing a gear mechanism in which gearsmesh with each other according to another embodiment of the presentinvention;

FIG. 21 is a schematic diagram showing a gear mechanism in which gearsmesh with each other according to still another embodiment of thepresent invention;

FIG. 22 is a cross-sectional view showing the construction of a dampingmechanism in a conventional gear mechanism;

FIG. 23 is a graph showing the relationship between the amount ofrelative rotation of gears and the resultant elastic force generated ina damping mechanism;

FIG. 24 is a cross-sectional view showing a damping mechanism of a gearmechanism as a modified example of the first embodiment of the presentinvention; and

FIG. 25 is a diagram modelling the gear mechanism of the modifiedexample of FIG. 24.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS First Embodiment

The first embodiment of the present invention will be described withreference to FIGS. 1 to 9. In the first embodiment, a gear mechanismconstructed according to the present invention is employed as a balancerapparatus of an in-line four-cylinder internal combustion engine.

First, the balancer apparatus that employs the gear mechanism of thepresent invention will be briefly described with reference to FIGS. 1and 2. FIG. 1 is a schematic view showing the structure of the balancerapparatus as viewed from one side thereof, and FIG. 2 is a schematicview showing the gear arrangement of the balancer apparatus.

As shown in these figures, the balancer apparatus includes a crankshaft20 as an output shaft of the engine that is supported by a cylinderblock 11 and a crank case 12 (shown in FIG. 1) of the engine, and firstand second balance shafts 30 and 40 disposed below the crankshaft 20 inparallel therewith.

Each balance shaft 30, 40 is supported by first and second radialbearings 15 and 16 formed by the crank case 12 and a housing. It is,however, to be noted that only the radial bearings 15 and 16 supportingthe first balance shaft 30 are illustrated in FIG. 1, and the radialbearings 15 and 16 for the first and second balance shafts 30, 40 arenot illustrated in FIG. 2. A pair of unbalance weights 33, 43 aremounted on each balance shaft 30, 40 such that the corresponding secondradial bearing 16 is interposed between the weights 33, 43.

A pair of balance weights 22 per cylinder, that is, eight balanceweights 22 in total, are mounted on the crankshaft 20. Moreover, a crankgear 21 that rotates as a unit with the crankshaft 20 is mounted on thecrankshaft 20 at a position adjacent to a middle balance weight 22 a asone of the balance weights 22.

The first balance shaft 30 is provided with a first driven gear 31. Thefirst driven gear 31 meshes with the crank gear 21, and is rotatablerelative to the first balance shaft 30. The first driven gear 31 has adiameter that is equal to the radius of the crank gear 21. Moreover, thefirst balance shaft 30 is provided with a counter gear 32 locatedadjacent to the first radial bearing 15. The counter gear 32 ispress-fitted on the first balance shaft 30, and is coupled to berotatable together with the first balance shaft 30. The first drivengear 31 is operatively coupled with the counter gear 32 via a dampingmechanism 50 that allows relative rotation between the gears 31, 32.

As shown in FIG. 2, the second balance shaft 40 is provided with asecond driven gear 41 located adjacent to the first radial bearing 15(not shown in FIG. 2). The second driven gear 41 meshes with the countergear 32, and is coupled to be rotatable together with the second balanceshaft 40.

At the respective ends of the balance shafts 30 and 40, thrust bearings35 and 45 for limiting axial movement of the respective balance shafts30 and 40 are disposed adjacent to the respective first radial bearings15. Each thrust bearing 35, 45 has a recess 35 a, 45 a formed in itsportion located opposite to the center of gravity of the unbalanceweights 33, 43 (i.e., the lower side of FIG. 2) with respect to thecentral axis of the corresponding balance shaft 30, 40. Similarly, eachof the counter gear 32 and the second driven gear 41 has a recess 32 a,41 a formed in its portion located opposite to the center of gravity ofthe unbalance weights 33, 43 (i.e., the lower side of FIG. 2) withrespect to the central axis of the corresponding balance shaft 30, 40.

With the recesses 32 a, 35 a, 41 a and 45 a thus formed, the respectivecenters of gravity of the counter gear 32, the second driven gear 41 andthe thrust bearings 35 and 45 are shifted or made eccentric to be on thesame side as that of the centers of gravity of the unbalance weights 33and 43. Accordingly, the counter gear 32, the second driven gear 41 andthe thrust bearings 35 and 45 perform substantially the same function asthe unbalance weights 33 and 43, when they rotate together with thebalance shafts 30 and 40. As a result, the size and weight of theunbalance weights 33 and 43 can be reduced by the amounts correspondingto the volume of the recesses 32 a, 35 a, 41 a and 45 a.

Moreover, as described above, the respective centers of gravity of thecounter gear 32, the second driven gear 41 and the thrust bearings 35and 45 are offset from the respective central axes of the balance shafts30 and 40. As the elements 32, 35, 41 and 45 rotate, therefore, thebalance shafts 30 and 40 are subjected to centrifugal force about therespective central axes of the balance shafts 30 and 40, at theirportions 30 a and 40 a supported by the corresponding first radialbearings 15.

Accordingly, the balance shafts 30 and 40 rotate with their supportedportions 30 a and 40 a being pressed against the respective innercircumferential surfaces of the first radial bearings 15 by thecentrifugal force. As a result, irregular or non-uniform vibration thatwould otherwise occur at the supported portions 30 a and 40 a can besuppressed when the balance shafts 30 and 40 rotate, whereby contactnoise that would be generated between each of the portion 30 a and 40 aand the inner surface of the corresponding first radial bearing 15 canbe reduced.

FIG. 5 schematically shows the relationship between the gears and theshafts. With the balancer apparatus of the engine constructed in theabove manner, rotational force is transmitted from the crankshaft 20 tothe first balance shaft 30, through the crank gear 21, the first drivengear 31, the damping mechanism 50 and the counter gear 32, and isfurther transmitted from the counter gear 32 to the second balance shaft40 through the second driven gear 41. In FIG. 5, reference characters“m1”, “m2” and “m3” denote the central axes of the crankshaft 20, thefirst balance shaft 30 and the second balance shaft 40, respectively.

Hereinafter, the structure of the damping mechanism 50 will be describedwith reference to FIGS. 3 and 4, each showing a cross section of thedamping mechanism 50 mounted on the first balance shaft 30. Morespecifically, FIG. 3 is a cross-sectional view taken along line 3—3 ofFIG. 4, and FIG. 4 is a cross-sectional view taken along line 4—4 ofFIG. 3.

As shown in FIG. 4, the first driven gear 31 includes an annular,radially inner portion 31 a that is disposed coaxially with the firstbalance shaft 30 so as to be rotatable relative to the first balanceshaft 30, and a radially outer portion 31 b that is disposed on theouter periphery of the radially inner portion 31 a so as to be rotatabletogether with the radially inner portion 31 a. The radially outerportion 31 b has teeth 31 c formed on its outer periphery. The teeth 31c of the radially outer portion 31 b mesh with teeth (not shown) formedon the outer periphery of the crank gear 21. Note that, in the presentembodiment, the teeth 31 c of the radially outer portion 31 b have thesame tooth width as that of the teeth of the crank gear 21. Moreover,the counter gear 32 has the same tooth width as that of the seconddriven gear 41 that meshes with the counter gear 32.

The radially inner portion 31 a of the first driven gear 31 is formed ofa metal such as iron, while its radially outer portion 31 b is formed ofa resin material consisting of a thermosetting resin, such as polyaminoamide or phenol, that is reinforced by aramid-fiber fabric. Similarly,at least the teeth of the second driven gear 41 are formed of theabove-described resin material. The crank gear 21 and the counter gear32 are both formed of a metal, such as iron. As shown in FIG. 5, thesegears 21, 31, 32 and 41 are helical gears formed with helical teeth.

The radially inner portion 31 a of the first driven gear 31 has a recess53 formed on the side opposite to that facing the counter gear 32. Therecess 53 is formed about the central axis of the first balance shaft 30so that the inner portion 31 a that radially defines the recess 53 hasan inside diameter that is larger than the outside diameter of the firstbalance shaft 30. With the first driven gear 31 engaging with the firstbalance shaft 30, therefore, an annular space is formed between theouter circumferential surface of the first balance shaft 30 and theinner circumferential surface of the radially inner portion 31 a (or theinner wall of the recess 53). A pair of annular friction dampers 54serving as damping members are disposed in the annular space.

Each friction damper 54 includes an elastic portion 54 a formed of anelastic material such as a rubber material, and a sliding portion 54 bformed of a metal and abutting on the inner wall surface of the recess53. The first driven gear 31 is always biased outwards in the radialdirection over the entire circumference of the first balance shaft 30,due to elastic force generated by the elastic portions 54 a of thefriction dampers 54.

Accordingly, when the first driven gear 31 rotates relative to thecounter gear 32, namely, when the first driven gear 31 rotates relativeto the first balance shaft 30, frictional force corresponding to themagnitude of the biasing force is generated between the sliding portions54 b and the inner wall surface of the recess 53. The frictional forceserves as damping force that acts against relative rotation between thefirst drive gear 31 and the counter gear 32.

The crank gear 21 and the first driven gear 31 are formed as helicalgears. Therefore, the first driven gear 31 can slightly move in theaxial direction of the first balance shaft 30 even with the crank gear21 and the first driven gear 31 meshing with each other. As a result,the first driven gear 31 may vibrate in the axial direction of the firstbalance shaft 30 due to fluctuations in rotational force or the like,and may be repeatedly brought into contact with the first balance shaft30, thereby possibly generating noise. The frictional force of thefriction dampers 54 also acts as damping force for damping the vibrationof the first driven gear 31.

The counter gear 32 has an annular recess 51 formed on the side facingthe first driven gear 31. The recess 51 is formed about the central axisof the first balance shaft 30 so as to surround the outer periphery ofthe first balance shaft 30. A plurality of engaging projections 52 (inthis example, four engaging projections as shown in FIG. 3) projectingtoward the first driven gear 31 are formed on an inner bottom surface 51a of the recess 51 at equal angular intervals about the central axis ofthe first balance shaft 30. These engaging projections 52 have agenerally rectangular shape in cross section. Moreover, a pair ofengaging holes 57 are formed in the inner bottom surface 51 a of therecess 51 at positions that interpose each of the engaging projections52 therebetween.

Furthermore, a plurality of stopper rubbers 55 (four stopper rubbers inthis example) that engage with the respective engaging projections 52and engaging holes 57 are provided in the recess 51 at equal angularintervals about the central axis of the first balance shaft 30. Thesestopper rubbers 55 have a generally trapezoidal shape in cross section.

Each stopper rubber 55 has an engaging recess 55 c that engages with thecorresponding engaging projection 52, and engaging pieces 55 d thatengage with the corresponding pair of engaging holes 57. Such engagementbetween the engaging projections 52 and the engaging recesses 55 c andengagement between the engaging pieces 55 d and the engaging holes 57limit movement of the stopper rubbers 55 in the circumferentialdirection within the recess 51. In the present embodiment, portions ofeach stopper rubber 55 that are located on both sides of thecorresponding engaging projection 52 while the stopper rubber 55 isengaged with the engaging projection 52 have the same length ordimension in the circumferential direction. In addition, the springconstant of each stopper rubber 55 is set or controlled so as not tocause excessive deformation leading a damage even when the maximumrotational force is applied from the crankshaft 20 to the balancerapparatus.

The radially inner portion 31 a of the first driven gear 31 has aplurality of projections 56 (four projections in this example) formed onthe side facing the counter, gear 32. These projections 56 projectingtoward the counter gear 32 are provided at equal angular intervals aboutthe central axis of the first balance shaft 30. More specifically, eachprojection 56 is spaced apart from the opposed ends of the adjacent twostopper rubbers 55 by predetermined angles θ1 and θ2, respectively.Although the predetermined angles θ1 and θ2 vary as the first drivengear 31 and the counter gear 32 rotate relative to each other, the sum(θ1+θ2) of these angles is a fixed value θmax (θmax=θ1+θ2).

Accordingly, the first driven gear 31 and the counter gear 32 can rotaterelative to each other within a predetermined rotational angle (=θmax)that is equal to the sum of the predetermined angles θ1 and θ2 (=θ1+θ2),before each projection 56 abuts on either end of the adjacent stopperrubbers 55. In other words, when the first driven gear 31 and thecounter gear 32 rotate relative to each other beyond the predeterminedrotational angle θmax, each projection 56 abuts on one end of thecorresponding stopper rubber 55. In the present embodiment, thepredetermined rotational angle θmax is set to “16°.”

Moreover, in the present embodiment, the number of teeth 31 c of thefirst driven gear 31 is set to an integral multiple of the number ofstopper rubbers 55. More specifically, the number of teeth 31 c “p” andthe number of stopper rubbers 55 “s” have the relationship defined bythe following equation (1):

p=n·s  (1)

where “s” and “n” are integers equal to or larger than 2.

In the balancer apparatus, the balance shafts 30 and 40 need be mountedso that respective rotational phases of the crankshaft 20 and eachbalance shaft 30, 40 have a predetermined relationship. Upon mounting ofthe balance shafts 30 and 40, therefore, respective positions of thebalance shafts 30 and 40 in the direction of rotation thereof areuniquely determined if the position of the crankshaft 20 in direction ofrotation thereof is determined.

When the first driven gear 31 is meshed with the crank gear 21 with therespective positions of the balance shafts 30 and 40 thus determined,the position of the first driven gear 31 in the direction of rotationthereof is also determined. When the counter gear 32 is subsequentlyassembled with the first balance shaft 30, therefore, it is required tolocate each projection 56 of the first driven gear 31 between theadjacent stopper rubbers 55 on the side of the counter gear 32, whilemeshing the counter gear 32 with the teeth of the second driven gear 41that is fixed to the second balance shaft 40.

If the number “p” of teeth 31 c of the first driven gear 31 and thenumber “s” of stop rubbers 55 are set to those having a relationship ofp≠n·s, unlike the present embodiment, there is only one way or manner inwhich the first driven gear 31 and the stopper rubbers 55 are positionedsuch that each projection 56 is located in a space between the adjacentstopper rubbers 55. Accordingly, if the angular position of the firstdrive gear 31 meshing with the crank gear 21 is not appropriate, eachprojection 56 cannot be located between the adjacent stopper rubbers 55when the counter gear 32 is brought into engagement with the firstbalance shaft 30 as described above. Thus, the first driven gear 31 needbe engaged with the crank gear 21 so as to be positioned at the onlyposition that permits subsequent mounting of the counter gear 32.

In the present embodiment, on the other hand, the teeth 31 c are formedat every (360/n·s)° in the circumferential direction of the first drivengear 31, and the projections 56 are formed at every (360/s)° in thecircumferential direction of the first driven gear 31. Accordingly, anangular interval between the projections 56 is an integral multiple ofthat between the teeth 31 c. Thus, the teeth 31 a of the first drivengear 31 are located at the same position no matter which one of theprojections 56 is located in a space between adjacent stopper rubbers55. As a result, the first driven gear 31 can be meshed with the crankgear 21 at a plurality of positions (four positions in the presentembodiment) while permitting subsequent mounting of the counter gear 32,thus assuring an increased degree of freedom with which the counter gear32 is mounted.

The balancer apparatus having the damping mechanism 50 as describedabove can be represented by a model as illustrated in FIG. 6.

First, when the rotational force F is transmitted from the crankshaft 20to the crank gear 21, the first driven gear 31 rotates relative to thecounter gear 32 with only the frictional force of the friction dampers54 acting as damping force (damping coefficient C2), until the angle ofrotation of the first driven gear 31 relative to the counter gear 32exceeds the predetermined rotational angle θmax, that is, until theprojections 56 abut on the respective stopper rubbers 55.

When the first driven gear 31 rotates relative to the counter gear 32beyond the predetermined rotational angle θmax, the projections 56 abuton the corresponding ends of the stopper rubbers 55, whereby the stopperrubbers 55 are elastically deformed in the circumferential directionthereof. As a result, the elastic force (spring constant K1) and dampingforce (damping coefficient C1) of the stopper rubbers 55 as a whole inaddition to the damping force of the friction dampers 54 act against therelative rotation between first driven shaft 31 and the counter gear 32.Thus, the rotational force F transmitted to the first driven gear 31 istransmitted to the first balance shaft 30 through the counter gear 32and is then transmitted from the counter gear 32 to the second balanceshaft 40 through the second driven gear 41, as the resultant force ofthe damping force of the friction dampers 54 and the elastic force anddamping force of the stopper rubbers 55.

Hereinafter, a vibration suppressing effect of the gear mechanismaccording to the present embodiment will be described with reference toFIGS. 7 to 9.

FIG. 7 is a graph showing experimental results regarding changes influctuations of angular acceleration of the first and second balanceshafts 30 and 40 with respect to the speed of revolution of the internalcombustion engine.

In FIG. 7, the solid line indicates a characteristic of the presentembodiment, and the one-dot chain line indicates a characteristic of acomparative example. In the comparative example, the damping mechanism50 is omitted, and the first driven gear 31 and the counter gear 32 aredirectly operatively coupled with each other so that the gears 31 and 32rotate together with each other.

As shown in FIG. 7, in the comparative example, the magnitude offluctuations in angular acceleration rapidly increases when therevolution speed of the engine exceeds 4,000 rpm. The reason for this isas follows: due to torsional resonance of the crankshaft 20, rotationalforce is transmitted from the crankshaft 20 to the balance shafts 30 and40 such that a sextic component of the fundamental frequency thereof isamplified, and the balancer apparatus resonates due to the sexticcomponent.

In the present embodiment, on the other hand, almost no resonancephenomenon due to the sextic component occurs even when the balancerapparatus receives rotational force whose sextic component of thefundamental frequency is amplified. It follows that the transmission ofthe sextic component is surely blocked or prevented.

FIG. 8 is a graph showing experimental results regarding changes influctuations of the angular velocity of the first and second balanceshafts 30 and 40 with respect to the speed of revolution of the internalcombustion engine. In FIG. 8, the vertical axis indicates thefluctuations in the angular velocity in logarithm.

In FIG. 8, the solid line indicates a characteristic of the presentembodiment in which the predetermined rotational angle θmax is set to“16°,” and the one-dot chain line indicates a characteristic of a firstcomparative example in which the rotational angle θmax is set to “8°.”The two-dot chain line of FIG. 8 indicates a characteristic of a secondcomparative example in which the rotational angle θmax is set to “0°.”Namely, in the second comparative example, opposite end faces of eachprojection 56 as viewed in the circumferential direction abut on thecorresponding end faces of the adjacent stopper rubbers 55 in anunloaded state.

It will be under stood from FIG. 8 that, in the second comparativeexample, there is a speed region within the range from 1,000 rpm to2,000 rpm of the revolution speed of the engine, in which region themagnitude of fluctuations in the angular velocity rapidly increases.This is because a secondary component of the fundamental frequency ofthe rotational force transmitted from the crankshaft 20 to the balanceshafts 30 and 40 causes resonance in the balancer apparatus.

In contrast, in the first comparative example, there is a region withinthe range from 1,000 rpm to 2,000 rpm in which the magnitude offluctuations in the angular velocity slightly increases, but the amountof the increase thereof is much smaller than that of the secondcomparative example.

The reason for this is as follows: since the rotational angle θmax isset to be larger than “0°,” the natural frequency of a vibration systemformed by the balancer apparatus is reduced, whereby the resonancephenomenon due to the secondary component can be suppressed in thenormal revolution speed range of the engine (>1,000 rpm). Another reasonis that large damping force is applied to the vibration system becauseof the frictional force that is generated at the friction dampers 54 asthe first driven gear 31 and the counter gear 32 rotate relative to eachother in the rotational phase range or angle θmax.

In the present embodiment in which the rotational angle θmax is set to“16°”, vibration-suppressing effects obtained by setting the rotationalangle θmax to be larger than “0°” and by using the friction dampers 54are further enhanced. In other words, while the revolution speed of theengine is in the range from 1,000 rpm to 2,000 rpm, the magnitude offluctuations in the angular velocity does not increase and theabove-described resonance phenomenon does not occur. The inventors haveconfirmed in a further detailed experiment that the resonance phenomenoncaused by the secondary component can be suppressed by setting therotational angle θmax to “1°” or larger.

FIG. 9 shows results of an experiment conducted to confirm thevibration-suppressing effect by the frictional force of the frictiondampers 54. Like FIG. 8, FIG. 9 is a graph showing changes in themagnitude of fluctuations in the angular velocity of the first andsecond balance shafts 30 and 40 with respect to the revolution speed ofthe internal combustion engine. In FIG. 9, the vertical axis representsfluctuations in the angular velocity in logarithm.

FIG. 9, the solid line indicates a characteristic of the presentembodiment, and the one-dot chain line indicates a characteristic of acomparative example. In the comparative example, the friction dampers 54are omitted, and the radially inner portion 31 a of the first drivengear 31 is supported on the first balance shaft 30 so as to be rotatablerelative to the first balance shaft 30.

As shown in FIG. 9, in the comparative example having no friction damper54, the magnitude of fluctuations in the angular velocity abruptlyreaches a peak when the revolution speed of the engine is in theneighborhood of 1,000 rpm. Namely, even if the rotational angle θmax isset to be larger than “0°,” an effect of suppressing the resonancephenomenon due to the secondary component can no longer be obtainedunless the damping force of an appropriate magnitude such as thefrictional force of the friction dampers 54 is applied when the firstdrive gear 31 and the counter gear 32 rotate relative to each otherwithin the rotational phase range or angle θmax.

It will be apparent from the above experimental results that, in thepresent embodiment, occurrence of the resonance phenomenon due to thesecondary and sextic components can be surely suppressed by setting therotational angle θmax to “1°” or larger, and generating the frictionalforce of the friction dampers 54 as damping force when the first drivengear 31 and the counter gear 32 rotate relative to each other withinthat rotational phase range or angle θmax. As a result, the springconstant of the stopper rubbers 55 can be set to be relatively large,and therefore each stopper rubber 55 is prevented from undergoingexcessive deformation that would cause a damage thereof.

As specifically described above, the gear mechanism of the presentembodiment provides the following effects or advantages.

(1) Where the rotational force transmitted from the crankshaft 20 to thebalance shafts 30 and 40 includes both a low-frequency component(secondary component) and a high-frequency component (sextic component)as vibration components, the resonance phenomenon due to the componentscan be favorably suppressed without causing any damage or functionaldeterioration of the damping mechanism 50.

(2) Since the frictional force generated at the friction dampers 54 actsas the damping force of the damping mechanism 50, the damping force canbe kept approximately constant without significantly changing dependingon the speed of relative rotation between the first driven gear 31 andthe first balance shaft 30. Accordingly, the capability of damping thelow-frequency component, such as the secondary component, as a vibrationcomponent of the rotational force can further be improved as comparedwith a structure that uses a so-called oil damper or the like as adamping member. Thus, occurrence of a resonance phenomenon due to thelow-frequency component can be further advantageously suppressed orprevented.

(3) The first driven gear 31 and the second driven gear 41 are formed asresin gears whose teeth are formed of a resin, and are thus able toabsorb impact that occurs at gear-meshing portions between the crankgear 21 and the first driven gear 31 and between the counter gear 32 andthe second driven gear 41, resulting in reduction in meshing noise.Furthermore, the fluctuations in the rotational force transmittedbetween the gears, in particular, its high-frequency component, can beadvantageously damped.

(4) With regard to the metal gears meshing with the resin gears (thedriven gears 31 and 41), that is, in the crank gear 21 and the countergear 32, the gear-meshing noise can be reduced even if the workingaccuracy of the tooth surfaces of the metal gears is somewhat low. It isthus possible to eliminate some process steps, such as shaving andpolishing the tooth surfaces of the metal gears, and backlash control byselection and adjustment of a shim commonly used for forming metalgears. Moreover, since the resin gears are respectively meshed with themetal gears, disadvantages such as thermal adhesion between the gearscan be avoided.

(5) By forming the recesses 32 a, 35 a, 41 a and 45 a, the respectivecenters of gravity of the counter gear 32, the second driven gear 41 andthe thrust bearings 35 and 45 are shifted away from the axes of thegears and bearings to be on the same side as that of the centers ofgravity of the unbalance weights 33 and 43. Accordingly, the countergear 32, the second driven gear 41 and the thrust bearings 35 and 45perform substantially the same function as that of the unbalance weights33 and 43, resulting in reduction in size and weight of the unbalanceweights 33 and 43.

(6) The centers of gravity of the counter gear 32, the second drivengear 41 and the thrust bearings 35 and 45 are respectively offset fromthe central axes of the balance shaft 30 and 40. Therefore, the balanceshafts 30 and 40 are subjected to the centrifugal force at theirportions supported by the respective first radial bearings 15.Accordingly, the balance shafts 30 and 40 rotate with the supportedportions thereof being pressed against the respective innercircumferential surfaces of the first radial bearings 15 by thecentrifugal force. As a result, irregular or non-uniform vibration atthe supported portions is suppressed, whereby contact noise generatedbetween each of the portions and the inner circumferential surface ofthe corresponding first radial bearing 15 can be reduced.

(7) The number of teeth 31 c of the first driven gear 31 is set to be anintegral multiple of the number of stopper rubbers 55. Therefore, thedegree of freedom in meshing the crank gear 21 with the first drivengear 31 and assembling the first driven gear 31 with the counter gear32, or the first balance shaft 30, via the damping mechanism 50, isincreased. Thus, the gear mechanism of the present embodiment can beassembled together with an improved efficiency.

(8) The resin gears (the first driven gear 31 and the second driven gear41) are formed of a thermosetting resin, such as polyamino amide orphenol, that is reinforced by aramid-fiber fabric. The resin gears thusformed exhibits excellent durability.

(9) The vibration of the first driven gear 31 in the axial direction ofthe first balance shaft 30 is damped by the frictional force of thefriction dampers 54, resulting in reduction or suppression of noisecaused by the vibration.

FIG. 24 and FIG. 25 show a modified example of the first embodiment inwhich elastic bodies 59, such as rubber dampers or metallic springs,having extremely small spring constant and damping coefficient areinterposed between corresponding end faces of the stopper rubbers 54 andprojections 56. Namely, the elastic bodies 59 substantially fillclearances L (corresponding to the above angles θ1, θ2) between thestopper rubbers 54 and the projections 56. Where K1 a and C1 a representthe spring constant and damping coefficient of the stopper rubbers 55,and K1 b and C1 b represent the spring constant and damping coefficientof the elastic bodies 59, as shown in FIG. 25, K1 b and C1 b are setsignificantly smaller than K1 a and C1 a, respectively. With thisarrangement, too, the resonance of the damping mechanism due to thesecondary vibration of the engine may be suppressed in the normalrevolution speed range of the engine, as in the case where therotational angle θmax is set to be larger than 0″. In addition, theabove arrangement may facilitate assembling of the driven gear 31 andcounter gear 32 with the damping mechanism 50 since the positions of thegears are automatically determined in the presence of the elastic bodies59. Furthermore, when the damping mechanism 50 of the modified exampleis mounted on a balance shaft, variations or shifts in the phase of theunbalance weights on the balance shaft may be eliminated, and thereforethe secondary vibration of the engine may be effectively suppressed.

Hereinafter, other embodiments of the present invention will bedescribed. As in the first embodiment, each of the following embodimentsis also applied to a balancer apparatus of a four-cylinder internalcombustion engine, and the basic structure thereof is the same as thatshown in FIGS. 1, 2 and 5. In the following, the difference between eachembodiment and the first embodiment will be mainly described. The sameelements as those described in the first embodiment are denoted by thesame reference numerals and characters, and description thereof will notbe provided.

Second Embodiment

First, a gear mechanism according to the second embodiment of thepresent invention will be described. FIG. 10 shows a specificconstruction of the gear mechanism according to the second embodiment

As shown in FIG. 10, the second embodiment is different from the firstembodiment in that a portion 55 a of each stopper rubber 55 that extendsfrom one side surface 52 a of the corresponding engaging projection 52in the direction opposite to the direction of rotation of the countergear 32 (hereinafter, such a portion 55 a will be referred to as“acceleration-side portion 55 a”) has a length different from that of aportion 55 b of each stopper rubber 55 that extends from the other sidesurface 52 b of the corresponding engaging projection 52 in thedirection of rotation of the counter gear 32 (hereinafter, such aportion 55 b is referred to as “deceleration-side portion 55 b”).

The deceleration-side portion 55 b is a portion on which thecorresponding projection 56 abuts when the rotational force transmittedfrom the crankshaft 20 decreases mainly during deceleration of theengine. In contrast, the acceleration-side portion 55 a is a portion onwhich the corresponding projection 56 abuts when the rotational forcetransmitted from the crankshaft 20 increases mainly during accelerationof the engine. More specifically, the acceleration-side portion 55 a isa portion on which the corresponding projection 56 abuts not only duringacceleration of the engine but also during a steady-state operation ofthe engine, i.e., when approximately constant rotational force istransmitted. Accordingly, the acceleration-side portions 55 a are morefrequently subjected to elastic deformation and are likely to undergo alarger amount of deformation, as compared with the deceleration-sideportions 55 b. Accordingly, the acceleration-side portions 55 a isrequired to be more durable than the deceleration-side portions 55 b.

Thus, in the present embodiment, the length Lb of the deceleration-sideportion 55 b is reduced, and the length La of the acceleration-sideportion 55 a is increased by an amount of reduction in the length Lb, sothat the lengths La and Lb of the portions 55 a and 55 b have arelationship of La>Lb. As a result, the volume of the acceleration-sideportion 55 a is larger than that of the deceleration-side portion 55 b,whereby the elastic deformation limit, i.e., the maximum permissibleamount of elastic deformation, of the acceleration-side portion 55 a islarger than that of the deceleration-side portion 55 b.

Accordingly, the acceleration-side portions 55 a can effectivelyalleviate and absorb impact that acts on these portions 55 a, and areallowed to elastically deform to a great extent without incurring anydamage thereof even when the rotational force from the crankshaft 20 israpidly increased.

On the other hand, since the deceleration-side portions 55 b are lessfrequently subjected to elastic deformation and have a less amount ofdeformation as compared with the acceleration-side portions 55 a, thelength Lb of the deceleration-side portions 55 b is set to be short.Therefore, the volume of the stopper rubber 55 is not unnecessarilyincreased as a result of increasing the length La of theacceleration-side portion 55 a.

Thus, according to the present embodiment, the following effects can beobtained in addition to the effects (1) to (9) as described in the firstembodiment.

(10) Since the acceleration-side portions 55 a of the stopper rubbers 55are allowed to elastically deform to a greater extent, the durability ofthe stopper rubbers 55 can be improved.

(11) Since the volume of the stopper rubber 55 is not increased to belarger than required, the rotational angle θmax can be kept large enoughto maintain the desired damping capability of the damping mechanism 50.

Third Embodiment

Hereinafter, a gear mechanism according to the third embodiment of thepresent invention will be described. The third embodiment is differentfrom the first embodiment in the point as follows. In the firstembodiment, the first driven gear 31 and the crank gear 21 have the sametooth width, and the counter gear 32 and the second driven gear 41 havethe same tooth width. In the third embodiment, however, the resin gearssuch as the first driven gear 31 and the second driven gear 41 have atooth width different from that of the metal gears such as the crankgear 21 and the counter gear 32 that mesh with the resin gears.

In general, the maximum rotational force transmitted between a pair ofgears meshing with each other is obtained, and the tooth width of eachof the gears is set so that the teeth will not be broken or damaged whenthey receive the maximum rotational force. In the case where a metalgear is meshed with a resin gear, therefore, the tooth width of themetal gear is determined in accordance with the tooth width of the resingear having a lower tooth strength.

As shown in FIG. 11A, it would be desirable if a resin gear Gr and ametal gear Gm mesh with each other in an accurately aligned manner inthe tooth-width direction of the gears. However, as shown in FIG. 11B or11C, the resin gear Gr may actually mesh with the metal gear Gm with theteeth of the gear Gr being displaced from those of the gear Gm in thedirection of their rotation axes.

Such displacement may be caused by an error in mounting the gears,vibration during rotation, and the like. Moreover, in the case where theresin gear and the metal gear are formed as helical gears, thrusts areapplied in different directions to the respective gears during rotationthereof, and the gears may be displaced or offset from each other to asignificant extent.

Such displacement in the tooth-width direction causes abutment betweenonly local portions of the metal gear and the resin gear. In such acase, a contact area between the respective teeth of the resin gear andthe metal gear is accordingly reduced, whereby a contact pressure isincreased. Accordingly, the resin gear having the lower wear resistanceand durability than the metal gear may be increasingly worn or damaged.

In the present embodiment, therefore, the first driven gear 31 (resingear) meshing with the crank gear 21 (metal gear) has a larger toothwidth than that of the crank gear 21. Similarly, the second driven gear41 (resin gear) also has a larger tooth width than that of the countergear 32 (metal gear).

By setting the respective tooth widths of the driven gears 31 and 41 inthe above manner, local abutment between the gears is avoided, and anotherwise possible increase in the contact pressure at the teeth of thedrive gears 31, 41 can be prevented even if the gears are displaced oroffset in the tooth-width direction.

FIG. 12 shows an experimental result regarding changes in the strengthof a toothed portion of a resin gear meshing with a metal gear when thetooth width Br of the resin gear is changed while the tooth width Bm ofthe metal gear is kept constant.

In FIG. 12, the horizontal axis indicates ratio Br/Bm (tooth-widthratio) of the tooth width Br of the resin gear to the tooth width Bm ofthe metal gear, and the vertical axis indicates the strength of thetoothed portion of the resin gear. In FIG. 12, the strength of thetoothed portion is represented as a relative value to a reference value“1.0” that is the strength when the tooth-width ratio Br/Bm is equal to“1.0.”

It will be understood from FIG. 12 that the strength of the toothedportion can be increased by setting the tooth-width ratio Br/Bm to behigher than “1.0.” In order to surely increase the strength of thetoothed portion of the resin gear, it is desirable to set thetooth-width ratio Br/Bm to “1.1” or larger. However, the strength of thetoothed portion is hardly increased once the tooth-width ratio Br/Bmexceeds “1.5.” Therefore, in order to prevent an increased size of theresin gear, it is desirable to set the tooth-width ratio Br/Bm to “1.5”or smaller.

In the present embodiment, the tooth width of the first driven gear 31is set to 1.1 times that of the crank gear 21, based on the relationshipbetween the tooth-width ratio Br/Bm and the strength of the toothedportion of the resin gear. Similarly, the tooth width of the seconddriven gear 41 is also set to 1.1 times that of the counter gear 32.

According to the present embodiment as described above, the followingeffect can be obtained in addition to the effects (1) to (9) asdescribed in the first embodiment.

(12) Even if the tooth positions of the first driven gear 31 and thecrank gear 21 or the tooth positions of the second driven gear 41 andthe counter gear 32 are displaced or shifted in the tooth-widthdirection, such displacement will not cause abutment of local portionsof these gears 31 and 21 (or 41 and 32). Accordingly, the driven gears31 and 41 do not suffer from wear due to the local abutment as describedabove, and are also free from damage and breakage.

Fourth Embodiment

Hereinafter, a gear mechanism according to the fourth embodiment of thepresent invention will be described.

Although a resin gear whose strength of the toothed portion is lowerthan that of a metal gear is used as the first driven gear 31 and thesecond driven gear 41, in the first embodiment, each of the gears 31 and41 ensures a sufficiently high degree of durability since the dampingmechanism 50 as described above favorably prevents the resonance fromoccurring in the gears. However, if the shift position isinappropriately changed on the transmission side of the engine, forexample, excessive rotational force that is not supposed to be receivedmay be transmitted from the crankshaft 20 to the balance shafts 30 and40, whereby the teeth of the driven gears 31 and 41 may be broken. Ifthe teeth of the driven gears 31 and 41 are broken, biting may occurbetween the crank gear 21 and the first driven gear 31 and between thecounter gear 32 and the second driven gear 41. Such biting may causeexcessive impact force to be applied to the crankshaft 20 or othermembers that rotate therewith, thereby possibly damaging the crankshaft20 and other members.

In the present embodiment, when excessively large rotational force istransmitted from the crankshaft 20 to the balance apparatus, therotational force from the crankshaft 20 is forcibly cut off or blocked.

More specifically, in the gear mechanism of the present embodiment, thebreakage strength as measured upon breakage (fracture) of each of theprojections 56 due to the elastic force of the stopper rubbers 55 is setto be smaller than the strength of the toothed portion of the firstdriven gear 31.

FIG. 13A is a schematic view illustrating the manner of measuring thestrength of the toothed portion of the first driven gear 31. FIG. 13B isa schematic view illustrating the manner of measuring the breakagestrength of each projection 56.

As shown in FIG. 13A, upon measuring the strength of the toothed portionof the first driven gear 31, the first driven gear 31 (the radiallyouter portion 31 b) is first fixed to a rotary shaft 200, and a lever201 is also fixed to the rotary shaft 200. Moreover, a stationarytoothed piece 203 whose teeth 202 have the same shape as those of thecrank gear 21 is meshed with the first driven gear 31. Then, a load isapplied to an end portion of the lever 102 in the direction of rotationthereof, so that the teeth 31 c of the first drive gear 31 engaging withthe toothed piece 203 are broken, and the load “fmax1” applied at thetime of the breakage is measured. The strength of the toothed portion ofthe first driven gear 31 is calculated as torque T1 (=fmax1·L1) that isobtained by multiplying the load “fmax1” by a length (or distance fromthe central axis of the rotary shaft 200 to the point that is subjectedto the load) L1 of the lever 201.

To measure the breakage strength of the projections 56, on the otherhand, the first driven gear 31 (the radially inner portion 31 a) isfirst fixed to a rotary shaft 204, and a lever 205 is also fixed to therotary shaft 204, as shown in FIG. 13B. Moreover, movement of oneprojection 56 in the rotation direction is limited by a stationary jig206. Then, a load is applied to an end portion of the lever 205 in thedirection of rotation thereof, so that the projection 56 is broken, andthe load “fmax2” applied at the time of breakage is measured. Thebreakage strength of the projection 56 is calculated as torque T2(=fmax2·L2·n) that is obtained by multiplying the load “fmax2” by alength L2 of the lever 205 and the number of projections 56 “n” (n=4 inthe present embodiment).

Thus, design values for defining the strength of the first driven gear31 and the projections, for example, the shapes of the teeth 31 c andthe projections 56, are suitably determined so that the strength T1 ofthe toothed portion of the first drive gear 31 and the breakage strengthT2 of the projections 56 have a relationship of T2<T1. Also, thestrength of the toothed portion of the second driven gear 41 is set sothat the second driven gear 41 has the same strength as that of thefirst driven gear 31.

In the gear mechanism of the present embodiment as described above, evenif such excessive rotational force that would break the teeth of thedriven gears 31 and 41 is transmitted from the crankshaft 20 to thebalance shafts 30 and 40, the projections 56 are broken before breakageof the gear teeth, whereby mechanical coupling between the first drivengear 31 and the counter gear 32 is forcibly cut off or eliminated. As aresult, the first driven gear 31 is no longer subjected to the inertialforce of the balance shafts 30 and 40 that is equal to or larger thanthe frictional force of the friction dampers 54. Accordingly, breakageof the first driven gear 31 can be surely avoided.

Similarly, the rotational force of the crankshaft 20 that is equal to orlarger than the frictional force of the friction dampers 54 is nottransmitted to the second driven gear 41. Therefore, breakage of thesecond driven gear 41 can also be surely avoided.

In the present embodiment as described above, the following effect canbe obtained in addition to the effects (1) to (9) as described in thefirst embodiment.

(13) Breakage of the toothed portions of the first and second drivengears 31 and 41 is avoided even when the excessive rotational force isreceived, and biting that would otherwise occur at gear meshing portionsincluding the first and second driven gears 31 and 41, and any problemcaused by the biting, can be prevented in advance.

Fifth Embodiment

Hereinafter, a gear mechanism according to the fifth embodiment of thepresent invention will be described. FIG. 14 shows a specific structureof the gear mechanism of the fifth embodiment. In the followingdescription, the stopper rubbers 55 are sequentially identified as afirst stopper rubber 551, a second stopper rubber 552, a third stopperrubber 553, and a fourth stopper rubber 554 in the circumferentialdirection of the counter gear 32.

One of the projections 56 that is located between an acceleration-sideportion 551 a of the first stopper rubber 551 and a deceleration-sideportion 552 b of the second stopper rubber 552 is identified as a firstprojection 561. The remaining projections 56 are sequentially identifiedas a second projection 562, a third projection 563, and a fourthprojection 564 in the circumferential direction of the counter gear 32.

As shown in FIG. 14, in the present embodiment, respectivecircumferential lengths La1, La2, La3 and La4 of the acceleration-sideportions 551 a to 554 a of the stopper rubbers 551 to 554 are set to bedifferent from each other. Similarly, respective circumferential lengthsLb1, Lb2, Lb3 and Lb4 of the deceleration-side portions 551 b to 554 bare set to be different from each other. As a result, in the presentembodiment, the projections 561 to 564 abut on the respective stopperrubbers 551 to 554 with different amounts of relative rotation betweenthe first driven gear 31 and the counter gear 32.

More specifically, the lengths La1 to La4 of the acceleration-sideportions 551 a to 554 a and the lengths Lb1 to Lb4 of thedeceleration-side portions 551 b to 554 b of the stopper rubbers 551 to554 are set to have the following relationship:

La1>La2>La3>La4  (2)

Lb1>Lb2>Lb3>Lb4  (3)

When the first driven gear 31 rotates relative to the counter gear 32 inthe same direction as that of rotation of the counter gear 32 while noneof the projections 561 to 564 is abutting on the corresponding stopperrubbers 551, 554, as shown in FIG. 14, the first projection 561initially abuts on the acceleration-side portion 551 a of the firststopper rubber 551. As the first driven gear 31 further rotates relativeto the counter gear 32, the second projection 562 abuts on theacceleration-side portion 552 a of the second stopper rubber 552. As thefirst driven gear 31 still further rotates relative to the counter gear32, the third projection 563 abuts on the acceleration-side portion 553a of the third stopper rubber 553, and finally, the fourth projection564 abuts on the acceleration-side portion 554 a of the fourth stopperrubber 554.

Similarly, when the first driven gear 31 rotates relative to the countergear 32 in the direction opposite to that of rotation of the countergear 32 while none of the projections 561 to 564 is abutting on thecorresponding stopper rubbers 551, 554, the fourth projection 564initially abuts on the deceleration-side portion 551 b of the firststopper rubber 551. As the first driven gear 31 further rotates relativeto the counter gear 32, the first projection 561 abuts on thedeceleration-side portion 552 b of the second stopper rubber 552. As thefirst driven gear 31 still further rotates relative to the counter gear32, the second projection 562 abuts on the deceleration-side portion 553b of the third stopper rubber 553, and finally, the third projection 563abuts on the deceleration-side portion 554 b of the fourth stopperrubber 554.

Accordingly, in the gear mechanism of the present embodiment, theoverall spring constant of the stopper rubbers 551 to 554 changes infour stages or steps depending upon the amount of relative rotationbetween the first driven gear 31 and the counter gear 32. Therefore, thestopper rubbers 551 to 554 have a more non-linear spring characteristic.As a result, the natural frequency of the vibration system formed by thebalancer apparatus is diversified into at least four frequencies, andtherefore the damping capability of the damping mechanism 50 can befurther enhanced. Moreover, since the projections 561 to 564 abut on therespective stopper rubbers 551 to 554 at different times, noise andvibration caused upon contact therebetween are alleviated.

According to the present embodiment as described above, the followingeffects can be obtained in addition to the effects (1) to (9) asdescribed in the first embodiment.

(14) The resonance phenomenon is not likely to occur at the naturalfrequency of the vibration system formed by the balancer apparatus.Thus, occurrence of the resonance phenomenon can be favorably prevented.

(15) Noise and vibration that would occur upon abutment of theprojections 561 to 564 on the respective stopper rubbers 551 to 554 canbe alleviated and thus reduced.

Sixth Embodiment

Hereinafter, a gear mechanism according to the sixth embodiment of thepresent invention will be described. FIGS. 15 and 16 show a specificstructure of the gear mechanism according to the sixth embodiment. Thesixth embodiment is different from the first embodiment only in thestructure of the damping mechanism 50.

More specifically, in the sixth embodiment, a damping structure 60constructed as described below is disposed between the first driven gear31 and the counter gear 32, as shown in FIGS. 15 and 16. FIG. 15 is across-sectional view taken along line 15—15 of FIG. 16, and FIG. 16 is across-sectional view taken along line 16—16 of FIG. 15.

As shown in FIG. 16, the radially inner portion 31 a of the first drivengear 31 has a projection 63 formed on the side facing the counter gear32. The projection 63 has a recess 63 a formed concentrically with thefirst balance shaft 30. A plurality of projections 63 b (threeprojections 63 b in this example) projecting further toward the countergear 32 are formed on the top face of the projection 63. A ring 34engages with the first balance shaft 30, to be located on the side ofthe first driven gear 31 with respect to the counter gear 32. The ring34 serves to limit axial movement of the first driven gear 31.

The counter gear 32 has an annular recess 61 formed on the side facingthe first driven gear 31. The annular recess 61 surrounds the outerperiphery of the first balance shaft 30. With the first driven gear 31engaging with the first balance shaft 30, an annular space is formed bythe outer circumferential surface of the projection 63 and an inner wallsurface 61 c that defines the recess 61. An annular friction damper 64serving as a damping member is disposed in this space. Like the frictiondamper 54 of the first embodiment, the friction damper 64 also includesa sliding portion formed of a metal, and an elastic portion formed of anelastic material such as a rubber material (both portions are not shownin the figures). The sliding portion abuts on the inner wall surface 61c of the recess 61, and the elastic portion abuts on the outercircumferential surface of the projection 63.

As shown in FIG. 15, a plurality of engaging grooves 62 (four engaginggrooves 62 in this example) each having a semi-circular cross-sectionare formed at predetermined intervals in an inner wall surface 61 a ofthe recess 61. These engaging grooves 62 are formed in the radialdirection of the first balance shaft 30. A plurality of metal coilsprings 65 (three coil springs 65 in this example) are disposed within aspace formed by the recesses 61 and 63 a. Moreover, a plurality oflimiting members 66 (three limiting members 66 in this example) forlimiting rotation of the respective coil springs 65 relative to thefirst balance shaft 30 as well as axial movement of the coil springs 65are provided within the space.

Each limiting member 66 has an engaging portion 66 a adapted forengaging with the corresponding engaging groove 62. By such engagementbetween the engaging portion 66 a and the corresponding engaging groove62, each limiting member 66 is fixed so as not to be rotatable relativeto the counter gear 32. The limiting members 66 and the coil springs 65are alternately disposed within the recess 61 such that opposite ends ofeach coil spring 65 abut on parts of the corresponding ends of theadjacent two limiting members 66, thereby preventing relative rotationof the coil springs 65.

Moreover, the projections 63 b are provided around the central axis ofthe first balance shaft 30 such that each projection 63 b is locatedbetween the adjacent coil springs 65 and is spaced apart from theopposed ends of the adjacent coil springs 65 by predetermined angles θ1and θ2, respectively. The predetermined angles θ1 and θ2 are changed asthe first driven gear 31 and the counter gear 32 rotate relative to eachother. However, the sum of the predetermined angles (θ1+θ2) is a fixedvalue (θmax=θ1+θ2).

Therefore, in the present embodiment as well, the first driven gear 31and the counter gear 32 can rotate relative to each other within apredetermined rotational phase range or angle (=θmax) that is equal tothe sum of the predetermined angles θ1 and θ2 (=θ1+θ2), before eachprojection 63 b abuts on either end of the adjacent coil springs 65. Inother words, when the first driven gear 31 and the counter gear 32rotate relative to each other beyond the predetermined rotational angleθmax, each projection 63 b abuts on the end of the corresponding coilspring 65. In the present embodiment, the predetermined rotational angleθmax is set to “10°”.

According to the structure of the present embodiment as described above,approximately the same effects as those of the first embodiment can beobtained.

It is to be understood that each of the illustrated embodiments may bemodified when appropriate in the manners as described below by way ofexample.

The manner in which the friction dampers 54 and 64 are mounted is notlimited to that described in each of the illustrated embodiments, butmay be suitably modified as long as frictional force of an appropriatemagnitude can be generated as the first balance shaft 30 and the firstdriven gear 31 rotate relative to each other.

In the sixth embodiment, the friction damper 64 is interposed betweenthe outer circumferential surface of the projection 63 formed on theradially inner portion 31 a of the first driven gear 31 and the innerwall surface of the recess 61 formed in the counter gear 32. It is,however, possible to modify the structure as shown in FIG. 17, such thatan annular support member 67 containing the coil springs 65 and having aprojection 67 b functioning equivalently to the projection 63 b is fixedto a side surface of the radially inner portion 31 a of the first drivengear 31 that faces the counter gear 32, and the friction damper 64 isdisposed between the support member 67 and the counter gear 32. In thestructure of FIG. 17, elements having the same or equivalent functionsas those of the elements described in the sixth embodiment are denotedby the same reference numerals and characters.

In the first through fifth embodiments, the gear mechanism may bemodified as shown in FIG. 18: an annular support member 58 is fixed to aside surface of the radially inner portion 31 a of the first driven gear31 that is located remote from the counter gear 32. In addition, anotherannular support member 59 is fixed to the first balance shaft 30 so asto face the support member 58. Then, a friction damper 74 constructedsimilarly to the friction damper 54 used in the first embodiment isdisposed between the support members 58 and 59.

While each of the friction dampers 54, 64 and 74 is constructed suchthat it includes a sliding portion and an elastic portion in theillustrated embodiments and the modification shown in FIG. 18, thesefriction dampers 54, 64 and 74 may alternatively be formed by a metalwave washer, a metal belleville spring, a metal wave spring, or thelike.

In the sixth embodiment, the coil springs 65 are used as elasticmembers. However, an arc spring having a circular-arc shape, a spiralspring, or the like may alternatively be used. In either case, amaterial forming the spring is not limited to a metal, but may be aresin material or a ceramic material.

While four stopper rubbers 55 or three coil springs 65 are provided aselastic members in the illustrated embodiments, any number of stopperrubbers and coil springs may be provided.

While the first driven gear 31 and the second driven gear 41 are resingears and the crank gear 21 and the counter gear 32 are metal gears inthe illustrated embodiments, other structures may be employed providedat least one of the gears meshing with each other is a resin gear. Forexample, the crank gear 21 and the counter gear 32 may be resin gears,or the crank gear 21 and the second driven gear 41 may be resin gears.

As schematically shown in FIG. 19A, in the gear mechanism in which aresin gear and a metal gear formed as helical gears mesh with eachother, a tooth-trace direction Dr of the resin gear Gr may be slightlychanged with a temperature rise (refer to the two-dot chain line in FIG.19A), even if the resin gear Gr is formed to have the tooth-tracedirection Dr parallel to a tooth-trace direction Dm of the metal gearGm.

When the tooth-trace direction Dr of the resin gear Gr is changed asdescribed above, an uneven contact pressure is produced at therespective meshing surfaces of the resin gear Gr and the metal gear Gm.As a result, the meshing surface of the resin gear Gr may be worn at aportion that is subjected to a high contact pressure.

Such a temperature rise of the resin gear Gr is mainly caused bytransmission of heat from the internal combustion engine. Therefore, thechange in the tooth-trace direction Dr of the resin gear Gr is settledas the engine temperature rises to a predetermined level to achieveequilibrium after starting of the engine. Thereafter, the tooth-tracedirection Dr is maintained approximately at the settled value.

Thus, as shown in FIG. 19B, it is desirable to preset the tooth-tracedirection Dm of the metal gear Gm in accordance with the resultanttooth-trace direction Dr of the resin gear Gr after the temperaturerise. More specifically, in each of the illustrated embodiments, thetooth-trace direction of the crank gear 21 may be made parallel to theresultant tooth-trace direction of the first driven gear 31 after thetemperature rise, and the tooth-trace direction of the counter gear 32may be made parallel to the resultant tooth-trace direction of thesecond driven gear 41 after the temperature rise. With this arrangement,local wear of the resin gear and an increase in meshing noise due tolocal abutment of the gears can be suppressed.

In each of the illustrated embodiments, as shown in FIG. 5, the gearmechanism is constructed such that the rotational force of thecrankshaft 20 is transmitted from the crank gear 21 to the first drivengear 31, and further transmitted from the first driven gear 31 to thecounter gear 32 through the damping mechanism 50 (60), and then,transmitted from the counter gear 32 to the balance shafts 30 and 40.The gear mechanism, however, may be constructed otherwise.

More specifically, as schematically shown in FIG. 20, the crank gear 21that is rotatable relative to the crankshaft 20, and a plate 25 that isrotatable together with the crankshaft 20 are provided on the crankshaft20. The crank gear 21 is operatively coupled with the plate 25 throughthe damping mechanism 50 (60). The first driven gear 31 and the countergear 32 that are rotatable together with the first balance shaft 30 areprovided on the first balance shaft 30, and the first driven gear 31 ismeshed with the crank gear 21.

With the gear mechanism thus constructed, the rotational force of thecrankshaft 20 is transmitted to the first balance shaft 30 through thedamping mechanism 50 (60), the plate 25, the crank gear 21 and the firstdriven gear 31, and is also transmitted to the second balance shaft 40through the counter gear 32 and the second driven gear 41.

The rotational force of the crankshaft 20 may alternatively betransmitted from the crank gear 21 to each of the balance shafts 30 and40 via separate transmission paths.

More specifically, as schematically shown in FIG. 21, the first drivengear 31 that is rotatable relative to the first balance shaft 30, and aplate 36 that is rotatable together with the first balance shaft 30 aremounted on the first balance shaft 30. The first driven gear 31 isoperatively coupled with the plate 36 via the damping mechanism 50 (60).Moreover, the second driven gear 41 that is rotatable relative to thesecond balance shaft 40, and a plate 44 that is rotatable together withthe second balance shaft 40 are mounted on the second balance shaft 40.The second gear 41 is operatively coupled with the plate 44 via anotherdamping mechanism 50 (60). In addition, the counter gear 32 that ismounted on a rotary shaft 37 different from the first balance shaft 30is meshed with the crank gear 21 and the second driven gear 41.

With the gear mechanism thus constructed, the rotational force of thecrankshaft 20 is transmitted from the crank gear 21 to the first balanceshaft 30 through the first driven gear 31, the damping mechanism 50 (60)provided on the first balance shaft 30, and the plate 36. Moreover, therotational force of the crankshaft 20 is also transmitted from the crankgear 21 to the second balance shaft 40 through the counter gear 32, thesecond driven gear 41, the damping mechanism 50 (60) provided on thesecond balance shaft 40, and the plate 44.

Even if the gear mechanism of each of the illustrated embodiments ismodified as shown in each of the above-mentioned figures, the modifiedembodiments provides the same effects or advantages as described in eachof the illustrated embodiments.

While the crank gear 21, the counter gear 32 and the driven gears 31 and41 are all formed as helical gears in the illustrated embodiments, thesegears may alternatively be formed as spur gears.

In the third embodiment, the tooth width of the resin gear (the firstdriven gear 31, the second driven gear 41) is set to 1.1 times that ofthe metal gear (the crank gear 21, the counter gear 32) that meshes withthe resin gear. However, the tooth width of the resin gear may be set toany value as long as the tooth width of the resin gear is larger thanthat of the metal gear. In order to avoid an increase in the size of theresin gear, it is desirable to set the tooth width Br of the resin gearto be in the range of Bm<Br<1.5×Bm (where Bm is the tooth width of themetal gear). In the case where a plurality of resin gears arerespectively meshed with a plurality of metal gears, respectivetooth-width ratios of the resin gears to the metal gears need not be thesame. For example, a tooth-width ratio of the first driven gear 31 tothe crank gear 21 may be different from that of the second driven gear41 to the counter gear 32.

In the fourth embodiment, the strength of the toothed portion of thesecond driven gear 41 is the same as that of the toothed portion of thefirst driven gear 31. However, the strength T3 of the toothed portion ofthe second driven gear 41 may be set to any value as long as thestrength T3 and the breakage strength T2 of the projections 56 have therelationship of T2<T3.

In the illustrated embodiments, two pairs of unbalance weights 33 and 43are provided on the respective balance shafts 30 and 40 so as tointerpose the corresponding second radial bearings 16 therebetween.However, the position and the number of unbalance weights 33 and 43 arenot limited to those of the above embodiments, and may be changed asdesired.

The respective structures of the gear mechanisms as shown in theillustrated embodiments may be combined when appropriate. For example,the gear mechanism according to the second embodiment may be furtherprovided with the structure of the third embodiment regarding the toothwidth of the resin gear, the structure of the fourth embodiment foravoiding breakage of the resin gear upon application of excessiverotational force, and the structure of the fifth embodiment fordispersing or diversifying the natural frequency. Moreover, the gearmechanism of the sixth embodiment may include the respective structuresas shown in the second through fifth embodiments.

While the gear mechanism is applied to the balancer apparatus of theinternal combustion engine in the illustrated embodiments, the presentinvention is not limited to this, but the gear mechanism mayalternatively be applied to another power transmitting system of theinternal combustion engine.

What is claimed is:
 1. A gear mechanism of a power transmitting systemcomprising first and second rotating members disposed coaxially witheach other, and a damping mechanism interposed therebetween, saiddamping mechanism including a damping member that generates dampingforce for limiting relative rotation between the first and secondrotating members, and at least one elastic member that elasticallydeforms mainly when an angle of relative rotation between the first andsecond rotating members exceeds a predetermined rotational angle, so asto apply elastic force onto the first and second rotating members in adirection opposite to that of the relative rotation, wherein saiddamping member comprises a friction damping member and said secondrotating member comprises a rotary shaft, and said friction dampingmember is disposed between said first rotating member and said rotaryshaft.
 2. A gear mechanism of a power transmitting system comprisingfirst and second rotating members disposed coaxially with each other,and a damping mechanism interposed therebetween, said damping mechanismincluding a damping member that generates damping force for limitingrelative rotation between the first and second rotating members, and atleast one elastic member that elastically deforms mainly when an angleof relative rotation between the first and second rotating membersexceeds a predetermined rotational angle, so as to apply elastic forceonto the first and second rotating members in a direction opposite tothat of the relative rotation, wherein said elastic member comprises amain deformation portion and a sub deformation portion, and the maindeformation portion has a greater deformation limit than the subdeformation portion.
 3. A gear mechanism according to claim 2, whereinsaid main deformation portion and said sub deformation portion comprisea rubber material, said main deformation portion having an elasticallydeformable portion whose volume is larger than a volume of anelastically deformable portion of said sub deformation portion.
 4. Agear mechanism according to claim 2, wherein said damping mechanismcomprises a plurality of elastic members, each of which is provided onone of said first and second rotating members, and a plurality ofabutting members corresponding to the respective elastic members, eachof which is provided on the other of said first and second rotatingmembers for abutting on the corresponding elastic members to causeelastic deformation thereof when the first and second rotating membersrotate relative to each other beyond the predetermined relativerotational angle, and the elastic members are located with respect tothe first and second rotating members such that different angles ofrelative rotation between the first and second rotating members areformed when the respective elastic members successively abut on thecorresponding abutting members.
 5. A gear mechanism of a powertransmitting system comprising first and second rotating membersdisposed coaxially with each other, and a damping mechanism interposedtherebetween, said damping mechanism including a damping member thatgenerates damping force for limiting relative rotation between the firstand second rotating members, and at least one elastic member thatelastically deforms mainly when an angle of relative rotation betweenthe first and second rotating members exceeds a predetermined rotationalangle, so as to apply elastic force onto the first and second rotatingmembers in a direction opposite to that of the relative rotation,wherein said damping mechanism comprises a plurality of said elasticmembers, each of which is provided on one of said first and secondrotating members, and a plurality of abutting members corresponding tothe respective elastic members, each of which is provided on the otherof said first and second rotating members for abutting on thecorresponding elastic members to cause elastic deformation thereof whenthe first and second rotating members rotate relative to each otherbeyond the predetermined relative rotational angle, and the elasticmembers are located with respect to the first and second rotatingmembers such that different angles of relative rotation between thefirst and second rotating members are formed when the respective elasticmembers successively abut on the corresponding abutting members.
 6. Agear mechanism of a power transmitting system comprising first andsecond rotating members disposed coaxially with each other, and adamping mechanism interposed therebetween, said damping mechanismincluding a damping member that generates damping force for limitingrelative rotation between the first and second rotating members, and atleast one elastic member that elastically deforms mainly when an angleof relative rotation between the first and second rotating membersexceeds a predetermined rotational angle, so as to apply elastic forceonto the first and second rotating members in a direction opposite tothat of the relative rotation, wherein said damping mechanism comprisesa plurality of said elastic members, each of which is provided on one ofsaid first and second rotating members, and a plurality of abuttingmembers corresponding to the respective elastic members, each of whichis provided on the other of said first and second rotating members forabutting on the corresponding elastic members to cause elasticdeformation thereof when the first and second rotating members rotaterelative to each other beyond the predetermined relative rotationalangle, and the predetermined relative rotational angle is defined by thesum of angles by which each said abutting member is spaced fromcorresponding end faces of said at least one elastic member which facethe abutting member, as viewed in a direction of rotation of saidrotating members.
 7. A gear mechanism of a power transmitting systemcomprising first and second rotating members disposed coaxially witheach other, and a damping mechanism interposed therebetween, saiddamping mechanism including a damping member that generates dampingforce for limiting relative rotation between the first and secondrotating members, and at least one elastic member that elasticallydeforms mainly when an angle of relative rotation between the first andsecond rotating members exceeds a predetermined rotational angle, so asto apply elastic force onto the first and second rotating members in adirection opposite to that of the relative rotation, wherein saiddamping mechanism comprises a plurality of first elastic members, eachof which is provided on one of said first and second rotating members, aplurality of abutting members corresponding to the respective firstelastic members, each of which is provided on the other of said firstand second rotating members, and a plurality of second elastic membersthat are interposed between corresponding end faces of the first elasticmembers and the abutting members; and each of the first elastic membershas a spring constant and a damping coefficient that are smaller than aspring constant and a damping coefficient of each of the second elasticmembers.
 8. A gear mechanism of a power transmitting system comprisingfirst and second rotating members disposed coaxially with each other,and a damping mechanism interposed therebetween, said damping mechanismincluding a damping member that generates damping force for limitingrelative rotation between the first and second rotating members, and atleast one elastic member that elastically deforms mainly when an angleof relative rotation between the first and second rotating membersexceeds a predetermined rotational angle, so as to apply elastic forceonto the first and second rotating members in a direction opposite tothat of the relative rotation, wherein at least one of said first andsecond rotating members comprises a gear meshing with a respectivecounter gear.
 9. A gear mechanism according to claim 8, wherein at leastone of said gears is a resin gear whose teeth comprise a resin material.10. A gear mechanism according to claim 9, wherein a respective gearmeshing with said resin gear is a metal gear whose teeth comprise ametal, and wherein said resin gear has a tooth width that is larger thana tooth width of the metal gear.
 11. A gear mechanism according to claim10, wherein said damping mechanism further comprises at least oneabutting member adapted for abutting on the corresponding elasticmembers to cause elastic deformation thereof when the first and secondrotating members rotate relative to each other beyond the predeterminedrelative rotational angle, and a strength of each said abutting memberas measured upon breakage of the abutting member due to elastic force ofthe corresponding elastic member acting thereon is set to be smallerthan a strength of a toothed portion of said gears.
 12. A gear mechanismaccording to claim 9, wherein said damping mechanism further comprisesat least one abutting member adapted for abutting on the correspondingelastic members to cause elastic deformation thereof when the first andsecond rotating members rotate relative to each other beyond thepredetermined relative rotational angle, and a strength of each saidabutting member as measured upon breakage of the abutting member due toelastic force of the corresponding elastic member acting thereon is setto be smaller than a strength of a toothed portion of said gears.
 13. Agear mechanism according to claim 8, wherein said damping mechanismfurther comprises at least one abutting member adapted for abutting onthe corresponding elastic members to cause elastic deformation thereofwhen the first and second rotating members rotate relative to each otherbeyond the predetermined relative rotational angle, and a strength ofeach said abutting member as measured upon breakage of the abuttingmember due to elastic force of the corresponding elastic member actingthereon is set to be smaller than a strength of a toothed portion ofsaid gears.
 14. A gear mechanism according to claim 8, wherein saiddamping mechanism comprises a plurality of said elastic members, each ofwhich is provided on one of said first and second rotating members, anda plurality of abutting members corresponding to the respective elasticmembers, each of which is provided on the other of the first and secondrotating members for abutting on the corresponding elastic members tocause elastic deformation thereof when said rotating members rotaterelative to each other beyond the predetermined relative rotationalangle, wherein the elastic members and said abutting members are locatedwith respect to said rotating members such that the elastic members andsaid abutting members are spaced from each other at equal intervals in adirection of rotation of said rotating members; and the first rotatingmember comprises a gear to be coupled by the gear mechanism, and thenumber of teeth of the gear is set to an integral multiple of the numberof the elastic members.
 15. An internal combustion engine comprising acrankshaft, a first balance shaft and a gear mechanism of a powertransmitting system comprising first and second rotating membersdisposed coaxially with each other, and a damping mechanism interposedtherebetween, said damping mechanism including a damping member thatgenerates damping force for limiting relative rotation between the firstand second rotating members, and at least one elastic member thatelastically deforms mainly when an angle of relative rotation betweenthe first and second rotating members exceeds a predetermined rotationalangle, so as to apply elastic force onto the first and second rotatingmembers in a direction opposite to that of the relative rotation,wherein said first balance shaft is driven by rotational torque of saidcrankshaft.
 16. An internal combustion engine according to claim 15,wherein said gear mechanism is arranged at said first balance shaft andcomprises a driven gear being disposed on said first balance shaft androtatable relative thereto, and said driven gear is driven by a crankgear fixedly secured on said crankshaft.
 17. An internal combustionengine according to claim 16, wherein a second balance shaft isoperatively coupled with said first balance shaft.
 18. An internalcombustion engine according to claim 16, wherein a second balance shaftis driven by said crankshaft via said crank gear, an intermediate gearbeing disposed on an intermediate shaft and meshing with said crankgear, a driven gear being disposed on said second balance shaft androtatable relative thereto and meshing with said intermediate gear, andan additional gear mechanism connecting said driven gear and said secondbalance shaft with each other.
 19. An internal combustion engineaccording to claim 15, wherein said gear mechanism is arranged at saidcrank shaft and comprises a crank gear disposed on said crankshaft androtatable relative thereto, and said crank gear drives a driven gearfixedly secured on said first balance shaft.
 20. An internal combustionengine according to claim 19, wherein a second balance shaft isoperatively coupled with said first balance shaft.